Armature and Rotor Balancing

Unbalance is defined as the uneven distribution of mass about the axis of rotation of a rotating body. Unbalance is the direct result of fixed or variable sources. Fixed source unbalance is caused by nonsymmetrical design or manufacture, while variable source unbalance is due to operational factors such as distortion or shifting of components by centripetal force. The result of the unbalance is a net centripetal force at the bearings, and this, in turn, causes vibrations in the system. Balancing is the procedure used to reduce the unbalance level to a level at which the resulting vibration is low enough that operation of the equipment is not impaired.

To define an unbalanced condition, four items must be identified. Unbalance magnitude U is the result of an unbalanced mass m located at a specific radius r. The location is defined first in terms of the angular location in degrees from a zero reference location, and finally by the axial location of the unbalance relative to the support bearings. In order to correct for unbalance in a rotor, you must identify these four parameters, represented graphically in Ill. 86. For this type of rotor, a single unbalance correction is adequate. For wider rotors, the resulting effects of multiple unbalances have to be resolved with two corrections at separate axial locations. This section explains these variations and provides guidelines to enable selection of appropriate production balancing methods.

ILL. 86 Location of unbalanced mass.

Unbalance magnitude U is the result of unbalanced mass m located at angle 45° on radius r.

Constituents of a Rotating Object

Shaft Axis. The shaft axis (centerline) of the rotor, shown in Ill. 87, is defined by the centerlines of the support bearings. The shaft axis may or may not contain the center of mass. Parts are designed to rotate about the shaft axis of a rotor, which is also the rotational axis.

Principal Axis. The principal axis always includes the center of gravity (or mass center) of the part. By definition, the mass of the rotor is evenly distributed about the principal axis. Unbalanced rotors attempt to rotate about the principal axis, but are restrained to rotate about the shaft axis by their support bearings.

The vibrations and bearing loads due to the unbalance are directly proportional to this misalignment. Ill. 87 presents an exaggerated illustration of how the mass of a rotor can be offset on the shaft axis. Note that the rotor is rotating around the shaft axis, but ideally, the center or principal axis is the true center of the rotor.

Center of Mass. The center of mass is generally considered to be at the geo metric center of an object. In fact, the center of mass, or inertia, is the location of symmetric mass distribution and is not the same as the location of the geometric center.

In an unbalanced rotor, the principal axis does not coincide with the shaft axis. When a rotor is balanced, the principal axis coincides with the shaft axis.

A rotor will always attempt to rotate about its principal axis but is con strained by the support bearings, which force it to rotate about the shaft axis.

Displacement of these two axes results in bearing loads and vibration which is directly proportional to the misalignment.

A rotating object is composed of components that have mass and flexibility and which absorb and dissipate energy when subjected to internal disturbances; the result is a unique pattern of motions called rotor response. The rotor response at operating speed directly affects the object's sensitivity to unbalance. Speeds close to resonant frequencies require much closer balance tolerances due to the increased rotor-response characteristics. The effects of speed and resonance are explained in more detail later.

Characteristics of Static and Dynamic Unbalance

Static Unbalance. Static unbalance is classified as single-plane unbalance, meaning that it’s represented by a single vector quantity and corrected by a single correction mass applied opposite to the unbalance location and in the axial plane of the mass center of the rotor. When there is no way to add (or remove) mass in the same plane as the unbalance, the correction can be split into two equal corrections at the ends of the rotor to achieve the same results. Static unbalance can be detected by non-rotational equipment and is characterized by the fact that the principal axis is displaced parallel to the shaft axis, as shown in Ill. 89.

ILL. 87 Misalignment of center of mass.

ILL. 88 Balanced rotor with even mass distribution.

Couple Unbalance. Couple unbalance is a condition in which the principle axis intersects with the shaft axis at the center of mass (which means there is no static unbalance) but is inclined at an angle to the bearing axis, as shown in Ill. 90. Couple unbalance (also called moment unbalance) has equal unbalances spaced 180° from each other, at opposite ends. Generally, couple unbalance is the result of each bearing plane not being precisely perpendicular to the rotational axis. When the rotor spins, it vibrates with a twisting motion. Couple unbalance cannot be corrected in one plane; two corrections are required.

Dynamic Unbalance. Dynamic unbalance follows the same condition as couple unbalance and can be measured only when the part is rotating. Dynamic unbalance is present when the central principal axis of inertia neither is parallel to nor inter sects the shaft axis at the center of mass. This can only be corrected in two or more planes. Dynamic unbalance is the general case of a combination of static and couple unbalance.

Quasi-Static Unbalance. In quasi-static unbalance, there is a specific combination of static and couple unbalance such that the angular position of one couple component coincides with the angular position of the static unbalance. If the axial location of the unbalance can be used for unbalance correction, then a single correction is possible. Otherwise, it must be treated the same as dynamic unbalance. In some production situations, correction can be made at a single plane close enough to the source of the unbalance to enable efficient balancing correction to within the required tolerance. This rare case of couple and static unbalance is shown in Ill. 91.

Multiplane balancing is sometimes required when there is insufficient material or space available to make all the required unbalance corrections. In this case, a pre balanced operation in one or two auxiliary planes precedes final balancing. In some applications where the rotors are flexible, multiplane balancing is used to minimize the rotor's internal bending stresses.

ILL. 89 Static unbalance: rotor with centered single correction.

ILL. 90 Couple unbalance: equal forces at opposite ends, 180° apart.

From a theoretical standpoint, the aim is always to achieve a zero-unbalance condition. In practice, the requirement is to reduce the unbalance to a point at which the unbalance forces have a negligible or nonharmful effect on part operation. Production situations require careful attention to the tolerances and correction procedures to achieve minimum cycle times with required service life and performance.

Effects of Increased Speed

The unbalance of a rotor does not change with speed but is solely due to its mass distribution about the bearing axis. The centripetal force generated by the unbalance increases with the square of the rotational speed change; therefore, the balance tolerance has to be determined for the maximum operating speed of the rotor.

The operating speed and bearing forces determine the bearing life, so the unbalance must be limited to achieve the required operating lifetime. The intensity of vibration (acceleration) must be limited to avoid noise and vibration apparent to the operator, which is proportional to the velocity of the vibration, which requires that the balancing tolerance be tightened as speed increases.

It’s normally not necessary to balance a rotor at operating speed. In fact, the aim should always be to balance at the lowest practical speed. A lower balancing speed is preferable for a number of reasons. Faster speeds require a longer time to ramp up and ramp down and also require stronger, stiffer tooling arrangements and tighter clamping of the rotor assembly, which in turn increases the risk of distortion or damage. Lower speeds assure safety for the operator and require less elaborate machine guards. However, the performance of the balancing machine transducers and instrumentation improves with increasing speed due to the greater signal levels from the increase in centripetal force, illustrated in Ill. 92.

ILL. 91 Quasi-static unbalance: combination of static and couple unbalance.

ILL. 92 Effects of centripetal force on mass.

The requirements of tooling design, balance tolerance, cycle time, and machine guarding all interact when determining the optimum balancing speed.

Centripetal Force

What doesn't move does not need balancing.

According to Newton's laws, a body in motion tends to move in a straight line unless acted on by an outside force that causes it to deviate from its normal course. In the case of rotating bodies, a particle on the outside travels on a course tangent to the rotation of the body. Centripetal force acts on the particle to accelerate it while the particle is attached to the rotor toward the center of rotation. If the rotor is balanced, then there is an apparent even distribution of mass. However, if there is an excess of mass concentrated at one location on the rotor (i.e., an unbalance), the rotating mass attempts to force the unbalanced portion in a tangential direction, thus causing vibration and fatigue of the components. Rotational motion is different from straight-line motion.

An automobile's acceleration stops when it reaches a constant speed. Turning a corner causes an instant acceleration even if the speed does not change; thus, acceleration is generated through the vehicle's tires. An unbalanced rotor experiences constant acceleration reacted through the bearings because the direction of the force is constantly changing. The velocity of the unbalanced mass is proportional to the speed, rpm, and the acceleration is proportional to the rate of change of the velocity, (rpm)^2 .

Components of centripetal force are the following:

_ Mass (volume)

_ Radius (distance)

_ Velocity (speed)

As an example, Ill. 93 shows a weight (mass) attached to a 3-ft string (radius) and twirled in a circle at 60 rpm (velocity). A greater mass produces a greater force, a longer string produces a greater force, and a faster speed produces a greater force proportional to the square of the speed.

Unbalance is independent of rotational speed and won’t increase or decrease when standing still. At zero speed, the unbalance has no effect on the rotor. However, if the rotor is rotated, the unbalance will exert centripetal forces, causing vibration that becomes more violent as the speed increases. Centripetal force increases proportionately to the square of the speed increase. At double the speed, centripetal force quadruples; triple the speed and the force is nine times as much.

Resonance

Every mechanical object has properties of mass, stiffness, and damping which deter mine its natural frequency of oscillation. Mass is the volume of the material times its density. Stiffness depends on the elasticity of the material. Damping is a measure of the ability of the system to dissipate vibratory energy.

The natural frequency is directly proportional to the stiffness and inversely proportional to the mass. This is the frequency at which the object will tend to self vibrate when rung by an impact.

Materials such as soft rubber have a high level of damping and a low stiffness and tend to absorb and dissipate vibration. Most hard materials have a higher stiffness and a lower level of damping. The damping factor determines the rate of energy loss to the surroundings. The damping factor is a nonlinear parameter and changes with speed. For a given structure, there is a frequency at which the damping factor approaches zero and therefore very little vibration energy is absorbed.

Resonance and critical speeds are frequencies that are governed by natural frequencies, damping, and vibratory forces. A resonance is a condition in a structure in which the frequency of a vibratory force, such as mass unbalance, is equal to a natural frequency of the system. If the vibratory force is caused by a rotating part, the resonance is called a critical speed.

A structure or object can be excited by one or more vibratory forces. Vibratory forces can be caused by various factors, including design, installation, manufacture, and wear, or the force can have a single constant frequency, as occurs with mass unbalance.

A rotational assembly with any finite unbalance acts as a vibration exciter and will produce a force as it’s rotated. This is called the excitation frequency. When the natural frequency and the excitation frequency coincide, a state of resonance is said to exist. As rotational speed approaches the resonant frequency, the effects of the force increases. At resonant frequency, vibration amplitudes can become very large.

If the rate of speed is close to the resonant frequency, a very low level of unbalance can still generate unacceptable vibration amplitudes.

As rotational speed reaches the resonant frequency, the support structure will vibrate directly with the exciting force (phase shift = 0°). As the speed increases nearer to resonance, the phase begins to shift until at resonance there is a 90° phase shift. As the rotational speed continues to increase, the phase continues to change until it reaches opposition (phase shift = 180°).

Balancing requires an exact knowledge of both the magnitude and the location of the unbalance, and so balancing speeds close to resonance are to be avoided. A small speed change will cause a large change in both the amount and the angle of the measured signal, and the results will be incorrect.

Sometimes equipment is designed to emphasize the resonant frequency. A tuning fork or piano string produces strong vibrations at the resonant frequency, which is beneficial; however, this is not the case with a stiff rotor, where the exact opposite condition is needed.

Vibrations that have large amplitudes can cause early fatigue failure. The energy expended by such vibrations causes’ significant power loss and speed reduction. In addition, noise levels from the vibration may be irritating to the operator as well as detrimental to the components surrounding the bearings.

It follows from this that as speeds and densities increase, keeping resonance away from the operating speed is a crucial part of the assembly designer's job. Ensuring that balancing speeds and tool design avoid resonance is a crucial part of the balancing-machine and tooling manufacturers' jobs.

ILL. 93 Components of centripetal force.

Correcting for Unbalance

Correction can be accomplished by adding, removing, or moving material. The physical properties of the rotating device must be analyzed to determine the best method of correction. Methods of correcting unbalance in a production environment are as follows:

_ Material removal by milling

_ Material removal by drilling

_ Material removal by cutting or clipping molded dimples

_ Material addition by weighted epoxies

_ Material addition by UV-cured epoxies

_ Adjustments to molds and dies

_ Relocation of the shaft axis (mass centering)

ILL. 94 Contour mill cut.

Whenever the unbalanced mass prohibits correction at a specified angle (as in an area where a correction cannot be made or the amount of added weight won’t all fit on a designated angle), correction can be made by vectoring adjacent angles or components to implement a correction. The total unbalance is split into two vectors for adjacent components. For high levels of unbalance, it may be necessary to vector more than two components and use more than two weights.

When designing a rotor, it’s important to estimate the maximum initial unbalance and to make allowance for adding sufficient correction mass at an appropriate location. The balance tolerance must be used to determine the amount of weight that can be added and the increments of weight in order to have enough resolution to bring the rotor into tolerance in one or two attempts.

A correction plane is determined by evaluating where mass can be added without affecting the mechanical operation of the rotor. Once a correction plane is selected, a correction method can be chosen. If weight is to be added, the weight mass and number of locations must be determined. Larger weight sizes permit more correction but limit the minimum value that can be achieved. If adhesive weights are to be used, a ridge, indentation, or roughened area is needed to prevent creep under loading.

Generally, weight removal is the preferred means of balancing small armatures and assemblies. Correction by drilling was the original preferred method; however this inefficient means of correction has been replaced almost exclusively by milling. Two of the most popular operations available on balancing machines are contour milling and component slot milling. Contour milling, shown in ill.94, involves the use of special milling cutters which are designed for the pro file or curve of the stack. Material is removed at the exact polar location of the unbalance. Contour milling allows for larger amounts of unbalance in one cut, which reduces the cycle time of the balancing process. This is ideally suited for manufacturers of armatures with the same stack diameter but different numbers of poles, or for balancing rotors with a skewed stack.

Component slot milling is used when accuracy is required. Grooves machined into the centers of adjacent pole tips are milled either by a V-mill or slot cutter. Ill. 95 shows this method of correction. The balancing machine calculates the unbalance vector using two or more components and determines how deep a cut should be made and whether the mill should transverse to adequately remove the unbalance mass. This type of milling should be used in cases in which the laminations are too thin to allow adequate material removal through profile con tour milling. This condition is exemplified by the edges of the pole tips becoming "feathered" because of a lack of material.

Weight removal offers several advantages over weight addition, especially for high-production applications: it’s faster, it can be more accurately con trolled, and it’s easier to automate.

However, hand milling has its advantages also. Whenever the armature has areas not easily accessed (such as fans), rotating components, or raised components, the armature or rotating assembly has to be corrected by hand. This is also true in low-production situations where automation is not justified.

To minimize cycle times it’s important to achieve tolerance in the smallest number of steps. The accuracy of the balancing machine must be adequate to give a precise readout of the required correction needed. Production can benefit greatly when the balancer instrumentation can convert the basic unbalance data into specific correction masses at specific locations so the operator has step-by-step guidance.

Types of Balancing Machines

Balancing machines are used to determine the unbalance of a rotor in terms of both the magnitude of the mass unbalance and its angular position, in one or more planes.

The system must support the rotor, spin it at a predetermined rpm, measure the forces that occur, and display the results in values useful for unbalance correction.

Automated machines may even index the rotor to the required angle for a correction and add correction weights or remove unbalance mass.

For production balancing, the machine is typically set up to provide some indication of whether the rotor is within tolerance. If the rotor is out of tolerance, precise readout of the angle and amount are required. Hard-bearing balancing machines typically respond directly to the rotor unbalance and are not affected by speed. They are also permanently calibrated.

Since the force that a given amount of unbalance exerts at a given speed is always the same, the output from the sensing elements attached to the balancing machine is directly proportional to the unbalance in the rotor. The variation of the sensing element with balancing speed is known, and compensation for speed variations is automatic. The output is not influenced by bearing mass, rotor mass, or inertia, so that a permanent relationship between unbalance and sensing element output can be established, and the display of unbalance amounts can be directly in terms of the required correction mass. The output is affected by looseness or by a lack of stiffness in the tooling, and it’s important that the tooling locates the housing (or bearing carrier) rigidly and without distortion.

The most popular method of obtaining a reference signal by which to determine the phase angle of the unbalance signal is through the use of an optical sensor that detects a reference mark representing zero degrees somewhere on the circumference of the rotor. A light source is reflected off the surface and picked up by the sensor. Through timing circuitry, an angle is established using the zero reference mark.

ILL. 95 V-mill or slot mill cutter.

The first balancing machines were developed around 1900; they were trial-and error devices consisting of a flexibly mounted set of bearings with some means of a drive. Some years later, mechanical indicators were developed. First the magnitude of vibration was indicated, then the high spot on the shaft was used as a crude reference for the angular position. Around 1940, electromechanical unbalance detection was developed, which was the predecessor to modern-day electronic balancing machines. Today, with the microprocessor, balancing machines have evolved into two sophisticated state-of-the-art, do-almost-anything machines categorized as soft bearing and hard-bearing balancing machines Soft-bearing balancing machines operate above the first critical speed, or resonance, of the machine/rotor system, as shown in Ill. 96.They determine unbalance in a rotor indirectly by measuring the displacement of the bearing bridge assembly that supports the rotor caused by the rotating part. Hence, the name soft, which actually refers to the relative stiffness of the bearing-support and spring system of the machine.

Every different part to be balanced on a soft-bearing machine causes a different displacement of the suspension system for a given amount of unbalance due to its weight, polar moments of inertia, and the like. This means that the machine has to be calibrated for each different workpiece type to be balanced. Soft-bearing machines were originally classified as trial-and-error balancing machines, because the first workpiece of each type had to be balanced to zero by trial and error before adding the known calibration weights. Modern electronics have simplified this procedure, but calibration of each rotor type is still required.

By contrast, hard-bearing machines operate below the critical speed of the machine/rotor system, as shown in Ill. 97. Hard-bearing machines directly mea sure the centripetal force created by the unbalance. Since a given amount of unbalance at any given speed causes the same force, regardless of rotor weight or shape, hard-bearing machines are classified as permanently calibrated. All that is required for setup is to input the geometric relationship between the bearing, or measuring, planes and the correction planes. This means that the unbalance in a rotor can be determined in the very first spin-up. No trial runs or calibration runs are required.

This obviously is a major advantage over soft machines, resulting in reduced balancing time, greater productivity, increased accuracy due to the elimination of the possibility of operator error, and increased simplicity, which eliminates the need for highly skilled or experienced operators.

ILL. 96 Soft-bearing balancing machines operate above resonance.

ILL. 97 Hard-bearing balancing machines operate below resonance.

The mechanical design and features of each type of machine also affect its physical properties. This can have an impact or limitation on the balancing capability of the system. As previously stated, the soft-bearing machine operates above the first critical speed. Therefore, when balancing a part, the part must be accelerated through the critical speed to bring it to the machine's balancing speed. Accelerating a rotor through the critical speed of the machine can cause several problems. First of all, it’s a safety hazard, because the rotor is much more susceptible to flying out of the machine due to the increased amplitude of the displacement. Second, it can dam age the machine, as well as the workpiece, by hitting the travel limits of the suspension. Third, the measuring system needs more time to stabilize and take reliable readings. The threat of these problems can be reduced by locking up the suspension system until the rotor reaches balancing speed. However, this creates its own set of problems by increasing cycle time due to an additional step for the operator and by introducing its own safety problem-the operator must approach the machine to unlock the suspension while a rotor is spinning.

The displacement measuring principal causes other problems as well. For example, external shocks to the machine caused by other equipment will cause excursions of the vibratory system. These excursions add to the displacement caused by the unbalance, but are totally indistinguishable from them. Similarly, certain rotor types, such as fans and blowers or even armatures with fans on them, will cause these excursions. Being incapable of separating the unbalance from the external influences, windage in this case, the operator will overcorrect the unbalance. This either will produce a bad part or, if the mistake is caught during an audit cycle, will require additional steps to bring the part into tolerance, rendering the operation inefficient.

Another workpiece-related problem that affects only soft-bearing machines is difficulties, or even damage, resulting from rotors with high levels of initial unbalance. When this condition occurs, the unbalance present in the part will cause the vibratory system to hit the end stops, possibly causing damage, but always preventing the machine from measuring the unbalance. The part must then be pre-balanced by some crude trial-and-error method so that the resulting displacement is within the measurable range of the machine.

The displacement measuring principle also affects the weight range of soft bearing machines. The bearings and related support components vibrate in unison with the rotor, adding to its mass. This parasitic mass reduces machine sensitivity by damping or reducing the displacement of the vibratory system. As the weight of the rotor approaches the weight of the parasitic mass of the machine, displacement of the bearing and, what is more important, of the transducers is reduced. This limits the low end of the weight capacity of the machine. This can be overcome by having additional sets of bearing supports with less parasitic mass, allowing smaller, lighter parts to be balanced. This raises the cost of the soft machine and adds yet another cumbersome operation, limiting flexibility.

Hard-bearing machines use the entire bearing support as part of the spring sys tem, eliminating parasitic mass altogether. Theoretically, hard-bearing machines have no lower weight limit. In actuality, however, they are limited by the physical size of the workpiece.

There is one further physical limitation of soft-bearing machines that is related to the type of workpiece to be balanced. This rotor type is classified as cantilevered or overhung. Balancing these parts requires special care on both types of machines. A cantilevered workpiece on a horizontal type of balance machine causes a negative load on the support furthest from the center of gravity (CG) of the part. On a hard machine, a counter-roller bearing must be used to hold the part securely in the machine. Measuring is not affected since the thrust bearing roller assembly is part of the vibratory system. On some types of soft machines, a downward load is required to ensure proper displacement of the vibratory system. This downward force can be created with an overhung part by special fixturing which holds the part so that the CG is between the bearing supports, creating a downward load on both supports. As in the case of using special roller carriages for small parts to extend the design limitations of the soft suspension, this arrangement allows balancing overhung parts, but at additional cost and time.

Balancing on a soft-bearing machine also requires repeatability of the balancing speed from run to run. The displacement of the suspension system on a soft machine is a function of the rotational speed. Therefore, it’s essential that on subsequent runs the balancing speed be identical to the speed at which the machine was calibrated. It’s clear from the preceding discussion that hard-bearing machines are the preferred choice for applications where frequent changeover and maximum flexibility are required. However, hard-bearing balancing machines have also become the norm in industries where the need for reliability and high-level accuracy outweigh all other considerations. These would include aircraft engines, turbo-machinery, and computer disk drives, to name only a few. In general, hard-bearing balancing machines have become the industry standard for all types of balancing applications.

Measuring System Sensitivity

The capability of a measuring system is determined by its sensitivity, repeatability, and accuracy. It’s a combination of the limitations of both the balancing machine itself and the intrinsic parameters of the rotor being measured.

The sensitivity of a balancing system is not the smallest display indication but the smallest amount that can be accurately measured. Sensitivity is typically measured in micro-inches.

The repeatability or precision of a system is the reproducibility of multiple measurements of the same part. For statistical evaluations, the repeatability test is carried out on a number of parts.

The accuracy of a system is the relation between the measured amount and the actual unbalance of the part. This is related to calibration of the system for a specific unbalance and linearity of the measurements with differing amounts of unbalance.

The smaller the difference, the more accurate the system. Consistent readings are an indication of repeatability, not accuracy.

The question is "What is the least amount of unbalance that can be detected?" Rotor operating characteristics play an important factor in determining the sensitivity of the machine. The initial unbalance, resonance, shaft concentricity and bearings of the rotor may limit the system. Bearing quality directly affects the noise level and therefore the signal-to-noise ratio of the unbalance signal. If the machine has to deal with a high initial unbalance, the measuring range has to be high to avoid overranging. The balancing machine's measuring range directly affects the amount of sensitivity. You can't measure ounces accurately if the range of the machine is in tons.

To find the limits for the system, a rotor and several removable calibration weights are needed. The rotor must have the capacity for mounting each of the calibration weights at several angular locations. The rotor should be balanced as closely as practical before measuring each of the calibration weights in each position. Plot ting each set of data should produce a sine wave, with the amplitude representing the remaining unbalance and the average representing the mass of the calibration weights. The average values indicate the linearity of the system, and the variations of the amplitude from the measured residual unbalance indicate the lower limits of machine measurement. Detailed analysis of accuracy is a subject in itself and is beyond the scope of this handbook.

Tooling

Balance tooling is the cradle or work-holding device used to securely hold a part during the unbalance measurement and correction processes. The tooling must sup port and locate the rotor to ensure that the unbalance forces (signal) are transferred to the unbalance detection transducers. If the balance tooling does not repeat, locate, and clamp the rotor assembly, the result will be errors in amount, angle, and axial distribution (for two-plane balance) of the unbalance measurement.

Balance tooling may include other functions, such as auto-indexing, remote angle, material addition/removal, and total indicated run-out (TIR), or contain the drive source for rotating the part. The tooling should be securely fastened to the measuring table or work supports on which the balancing operation is to be per formed.

Typically, the armature will want to creep to one side or another while spinning.

A stop or end thrust positioned so the armature shaft rides against it will maintain the rotating armature in a set position. When adjusted correctly, the end thrust won’t affect the unbalance measurement of the rotor. Misalignment may cause restriction in the free spinning of the armature and/or bouncing in the cradle.

To ensure precision in any balancing operation, the rotor must be properly positioned with respect to the work supports and correction planes. This is called referencing. To ensure the desired accuracy, the operator must make sure the part is precisely located and rigidly supported. To make the rotor easier to load and unload, adjust the end thrust to allow the rotor to ride against it but also to compliment the pole sensors. A properly adjusted balancing cradle will ensure correct plane location, short load and unload times, and simple operation. Achieving these often contradictory requirements takes careful attention to detail.

The most advanced and accurate balancing machine is worthless if the tooling is not adequate. A good reference for accurate unbalance measurements starts with perfected tooling.